Continuously variable belt drive system

ABSTRACT

A continuously variable belt drive system including a driving pulley assembly, a driven pulley assembly, and a V-shaped belt engaged to transfer rotary power therebetween is disclosed herein. The driving pulley assembly includes an idler pulley portion. As the belt speed decreases, the belt migrates downward within the driving pulley assembly until it falls within the idler pulley portions, thereby isolating the belt from the rotation of the driving pulley assembly. The driving pulley assembly also includes a belt tensioning mechanism that maintains proper belt tension at all speeds and drive ratios. The tensioning mechanism relies upon a pivotably mounted centrifugal weight arm to provide tensioning as a function of speed. A compression spring provides axial pressure to maintain belt tension over all drive ratios. The spring, acts against a curved cam edge on a pivoting lever arm, which arm transfers an axial force to the rear sheave of the driven pulley assembly.

CROSS-REFERENCE TO RELATED APPLICATIONS

This is a division of U.S. patent application Ser. No. 09/973,373, filedon Oct. 9, 2001 and issued as U.S. Pat. No. 6,962,543, which is adivision of U.S. patent application Ser. No. 09/405,188, filed on Sep.24, 1999 and issued as U.S. Pat. No. 6,406,390

BACKGROUND OF THE INVENTION

The present invention concerns a continuously variable transfer driveassembly or transmission mechanism, such as the type suited for use inautomotive applications to drive accessory devices. More particularly,the invention relates to a mechanically adjustable belt-type pulleysystem.

Automotive vehicles include a cooling system to dissipate heat developedby the vehicle power plant, such as an internal combustion engine. In atypical automotive vehicle, the lubrication system provides some coolingfunction as hot lubricant is pumped away from the engine. However, thebulk of the cooling requirements for the automotive vehicle isaccomplished by air flowing through the engine compartment and across aradiator. Coolant flowing around the power plant extracts heat from theengine, which heat is subsequently dissipated through the vehicleradiator.

In automotive vehicles, the engine compartment is designed to permitflow of ambient air through the compartment and past the radiator. Inmost vehicles, a cooling fan is provided that increases the flow of airacross the radiator. In some vehicle installations, the fan is driven byan electric motor that is independent of the vehicle engine. For smallerpassenger cars, the electric motor approach can satisfy the coolingneeds for the vehicle. However, unlike passenger cars, heavy truckscannot use electric motors to drive the cooling fan. For a typical heavytruck, the cooling fan would require up to 50 horsepower to cool theengine, which translates to unreasonably high electrical powerrequirements.

In a typical automotive installation, whether light passenger or heavytruck, the cooling fan is driven by the vehicle engine. In one typicalinstallation shown in FIG. 1, an engine 10 provides power through adrive shaft 11 to a transmission 12. Power to the driven wheels isaccomplished through a differential 14. In addition to providing motivepower, the engine 10 is also coupled to a transfer drive assembly 15.This assembly 15 provides power directly to a cooling fan 16 that ispreferably situated adjacent the vehicle radiator 17.

A wide range of technologies is available to transmit power from theengine 10 to the rotating cooling fan 16. For instance, some transferdrive assemblies 15 are in the nature of on/off clutches. The clutchesutilize a friction material to engage the fan when the clutch isactuated. A belt between an output shaft of the engine and the clutchprovides rotational input to the clutch in relation to the engine speed.In another drive assembly, a viscous fan drive relies upon the shearingof viscous fluid within a labyrinth between input and output members ofthe drive. The engagement of the drive is controlled by the amount offluid allowed into the labyrinth. Viscous drives suffer from manydeficiencies. For instance, drives of this type are inherentlyinefficient because a great amount of energy is lost in heating theviscous fluid. For many viscous drives, this parasitic power loss can beas high as five horsepower.

Another difficultly experienced by viscous fluid fan drives is known as“morning sickness.” When the vehicle is started cold, the fluid in thefan drive is more viscous than under normal operating conditions. Thishigher viscosity causes the drive to turn the cooling fan at full speed,which causes the cooling system to operate at maximum capacity during atime when the vehicle engine needs to be warming up. A further problemwith viscous fan drives is that they require a residual speed even whenfully disengaged. This residual speed is usually in excess of 400 r.p.m.and is necessary to allow enough fluid circulation within the drivelabyrinth for the drive to re-engage on demand.

The most prevalent transfer drive systems for a vehicle cooling systemrely upon a continuous belt to transfer rotational energy from thevehicle engine to the cooling fan. In the simplest case, one pulley isconnected to an output shaft of the engine and another pulley isconnected directly to the cooling fan. In this simple case, the speed ofthe cooling fan is directly tied to the engine, varying only as afunction of the fixed diameters of the two pulleys. Typically, the ratioof these diameters generates a speed ratio greater than 1:1—i.e., thefan pulley rotates faster than the engine pulley.

One problem exhibited by fixed pulley fan drives is that the fan speedis limited to the fixed ratio relative to the engine input speed. Formost vehicles, and particularly most heavy trucks, the maximum coolingair flow requirements occur at the engine peak torque operatingcondition, which is usually at lower engine speeds. Thus, in order toachieve the proper cooling flow rates, the cooling fan must be sized toprovide adequate cooling at the lower engine speeds. The power generatedby a fan is related to the cube of its speed. Thus, a fan sized to coolan engine at a lower speed, such as 1200 r.p.m., is grossly oversized athigher engine operating speeds, such as a typical rated speed of 2100r.p.m. From a cooling standpoint, the significantly greater coolingpower provided at higher speeds is not detrimental. However, thisover-sizing of the fan equates to wasted power when the engine is notoperating at its peak torque condition. For example, a typical 32-inchcooling fan operating at an engine rated speed of 2100 r.p.m., drawsapproximately 45 horsepower. Of this 45 horsepower, only a fraction, inthe range of 10 horsepower, is actually necessary to meet the engines'cooling requirements at this speed.

In order to address the varying cooling needs throughout an entireengine operating range, various cooling systems have been developed. Forinstance, in one type of system, the blades of the fan are rotated toprovide variable flow rates. In another application, the shapes of thefan blades themselves are altered to increase or decrease the flow rateat a constant fan rotational speed.

One approach to solving the problem of varying cooling needs in anautomotive setting has been the continuously variable transmission (CVT)or variable transfer drive assembly. In its most fundamental design, theCVT utilizes a continuous belt having a V-shaped cross section. The beltis configured to engage conical friction surfaces of opposing pulleysheaves. The continuously variable feature of the CVT is accomplished bychanging the distance between the sheaves of a particular pulley. As thesheaves are moved apart, the V-shaped belt moves radially inward to alower radius of rotation or pitch. As the sheaves are moved together,the conical surfaces push the V-shaped belt radially outward so that thebelt is riding at a larger diameter. The typical CVT is also sometimesreferred to as an infinitely variable transmission in that the V-beltcan be situated at an infinite range of radii depending upon thedistance between the conical pulley sheaves.

Much of the development work with respect CVT's has been in providing acontinuously variable transmission between a vehicle engine and itsdrive wheels. In a few instances, CVT's have been applied as anaccessory drive. For example, NTN Corporation has developed a rubberbelt CVT system that provides a constant accessory drive speedregardless of engine speed. The system using two spring-loadedadjustable pulleys, each having centrifugal weighs that compensate forchanges in engine speed. In this system, as the engine speed increases,the centrifugal weights translate radially outward to exert a force onone sheave pushing it toward an opposing sheave. This change in diameterof the sheave maintains a fixed rotational speed, even as the enginespeed increases, by altering the ratio of pulley diameters. This fixedspeed is used to maintain a constant alternator speed.

Ideally, a transfer drive assembly, such as assembly 15 shown in FIG. 1,would turn the cooling fan only as fast as is necessary to maintain anoptimal engine temperature. Controlling the cooling fan speed conservespower and improves the engine's overall efficiency. In addition, thetransfer drive assembly should have the ability to turn the fan fasterat lower engine speeds than at higher engine speeds, because the coolingrequirements for the engine are greater during operation at low speedand high torque.

Thus far, no accessory drive assemblies are known that are capable ofachieving all of these features. Although the continuously variabletransmission has been beneficial in operation of cooling fans, thetypical CVT cannot accomplish all of these particular factors.

SUMMARY OF THE INVENTION

The present invention contemplates a continuously variable belt pulleytransfer assembly that addresses these prior deficiencies. In oneembodiment, the transfer assembly includes a driving pulley assembly anda driven pulley assembly, with a continuous belt transferring rotarymotion therebetween. The pulleys are each formed by forward and rearsheaves that define opposing conical surfaces. The drive ratio betweenthe pulleys is determined by the position of the V-shaped belt betweenthe conical surfaces of the sheaves.

In one feature of the invention, one pulley assembly, preferably thedriving assembly, includes a belt tensioning mechanism that maintainsproper belt tension at any speed and pulley drive ratio. The mechanismcan include a weight arm that is pivotably mounted to a floating sleeve.The forward and rear sheaves forming the driving pulley are mounted tothe floating sleeve for rotation with the sleeve. The sleeve is splinedto a rotating drive shaft so the sleeve can slide freely along the driveaxis while rotational motion is transmitted to the sleeve. The floatingsleeve allows the driving pulley to align itself with the driven pulleywhen the driven pulley adjusts the drive ratio.

Rotation of the floating sleeve causes the weight arm to swing radiallyoutward due to centrifugal effects. The weight arm bears against aroller mounted on the rear sheave, thereby providing an axial force topush the rear sheave toward the relatively stationary forward sheave. Asthe floating sleeve and driving pulley rotate faster, the axial forcegenerated by centrifugal movement of the weight arm increases.

In another aspect of the tensioning mechanism, a spring and lever armconfiguration is used to maintain proper belt tension as the drive ratiochanges. The mechanism uses a spring plate tending to push the rearsheave toward the forward sheave. When the rear sheave is in itsforward-most position, a compression spring associated with the springplate is only slightly depressed so its axial force is minimal. Thepresent invention contemplates a lever arm disposed between thecompression spring and the rear sheave that helps maintain adequateaxial force even when the spring is at its minimum compression. Thelever arm is pivotably mounted to the floating sleeve and includes aroller at its free end that bears against the rear sheave. Thecompression springs are retained between the floating sleeve and aspring plate that is free to slide axially relative to the drivingpulley. The spring plate includes a roller that contacts a cam edge ofthe lever arm. Spring force is thus transmitted through the spring plateroller, to the lever arm and eventually to the rear sheave via anotherroller. The cam edge of the lever arm has a curvature that is calibratedto maintain the necessary axial force at all positions of the rearsheave, including its forward-most position.

In yet another feature of the invention, one of the pulleys, againpreferably the driving pulley, includes a disengagement mechanism thatisolates the belt from the rotation of the pulley. In one embodiment,the disengagement mechanism includes an idler pulley portion between theforward and rear sheaves of the driving pulley. The idler pulley portiondefines conical surfaces that transition into the conical surfaces ofthe primary pulley sheaves. The idler pulley portions are isolated fromthe forward and rear sheaves by bearings. As the belt sinks lower intothe pulley groove it eventually contacts the idler pulley portions. Atthis point, the belt is no longer in contact with the driving pulleysheaves, so rotation of the driving pulley is not translated to rotationof the belt.

The invention also contemplates improvements to a driven pulley member.The driven member includes a ratio adjustment mechanism that utilizes anelectric motor and gear arrangement to vary the distance of the rearsheave relative to the forward sheave of the pulley. An actuation screwis provided that can be threaded into and out of a split nut byoperation of the electric motor. As the actuation screw is threaded intothe split nut, it advances along the axis of the driven pulley assembly.As the screw advances it applies pressure through intermediatecomponents on the rear sheave, pushing it axially toward the forwardsheave. Conversely, as the actuation screw is unthreaded from the splitnut, the axial pressure on the rear sheave is relieved and the sheavemoves away from the forward sheave.

The invention further contemplates a fail-safe feature that restores thedriven pulley assembly to a predetermined drive ratio in the event of afailure of power to the electric motor. In one aspect, this featurerelies upon engagement fingers to hold the separable components of thesplit nut together to maintain the threaded engagement with theactuation screw. Once the components of the split nut are separated, theinternal threads of the nut are disrupted and the threaded engagementwith the actuation screw is terminated. In one embodiment, a solenoidholds the engagement fingers in contact with the split nut components.When power to the solenoid is interrupted, the solenoid can no longerhold the engagement fingers in position. A return spring can then pushthe fingers back, allowing the portions of the split nut to expandapart.

In accordance with certain features of the invention, once the split nutis disrupted, the actuation screw is driven forward by operation of alarge compression spring. As the actuation screw is propelled forward,it causes the rear sheave to be pushed forward until the sheave reachesa predetermined drive ratio position.

It is one object of the invention to provide a continuously variabletransfer system that provides mechanical adjustment of the drive ratioof the system. A further object is to provide such a system thatmaintains sufficient tension in the belt at all speeds and drive ratios.

A further object of the invention is accomplished by features thatrestore the transfer system to a predetermined drive ratio on theoccurrence of particular failures. Another object is to provide atransfer system that can achieve a wide range of drive ratios. Yetanother object achieved by the invention is to provide means fordisengaging the continuous belt from rotation under establishedconditions.

These and other objects, as well as several benefits of the inventioncan be readily discerned from the following written description of theinvention, as illustrated by the accompanying figures.

DESCRIPTION OF THE FIGURES

FIG. 1 is schematic representation of an engine, transmission andcooling system.

FIG. 2 is a block representation of one type of transfer drive assemblyutilizing a continuous belt and rotating pulley according to a preferredembodiment of the invention.

FIG. 3 is an enlarged side cross-sectional view of the driving member ofthe transfer drive assembly depicted in FIG. 2.

FIG. 4 is a side cross-sectional view of a forward pulley sheave of thedriving member assembly depicted in FIG. 3.

FIG. 5 is a side cross-sectional view of a rear pulley sheave of thedriving member assembly shown in FIG. 3.

FIG. 6 is an end elevational view of the rear sheave shown in FIG. 5.

FIG. 7 is an end elevational view of a floating sleeve used in thedriving member assembly shown in FIG. 3.

FIG. 8 is a side cross-sectional view of the floating sheave depicted inFIG. 7.

FIG. 9 is an end elevational view of a spring-plate used in the drivingmember assembly shown in FIG. 3.

FIG. 10 is a side elevational view of the spring-plate shown in FIG. 9.

FIGS. 11 and 12 are side partial cross-sectional representations of thedriving member assembly shown with the pulley sheaves in twoorientations.

FIG. 13 is a side cross-sectional view of a further embodiment of adriving member assembly for use as part of the transfer drive assemblyshown in FIG. 2.

FIG. 14 is a side cross-sectional view of a driven member assembly foruse with the transfer drive assembly shown in FIG. 2.

FIG. 15 is an end elevational view of the driven member assembly shownin FIG. 14.

FIG. 16 is an end elevational view of a rear sheave of the driven memberassembly shown in FIG. 14.

FIG. 17 is an end elevational view of bearing pressure plate used in thedriven member assembly shown in FIG. 14.

FIG. 18 is an end cross-sectional view of a support shaft used in thedriven member assembly shown in FIG. 14.

FIG. 19 is an end elevational view of a split nut used with the drivenmember assembly shown in FIG. 14.

FIG. 20 is a side elevational view of the split nut shown in FIG. 19.

FIG. 21 is an end elevational view of a retainer for the split nut foruse in the driven member assembly shown in FIG. 14.

FIG. 22 is a side partial cross-sectional view of an alternativeembodiment of a driven member assembly for use with a transfer driveassembly as shown in FIG. 2.

FIG. 23 is a side cross-sectional view of a further alternativeembodiment of a driving member assembly for use in the transfer driveassembly depicted in FIG. 2.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

For the purposes of promoting an understanding of the principles of theinvention, reference will now be made to the embodiments illustrated inthe drawings and specific language will be used to describe the same. Itwill nevertheless be understood that no limitation of the scope of theinvention is thereby intended. The invention includes any alternationsand further modifications in the illustrated devices and describedmethods and further applications of the principles of the inventionwhich would normally occur to one skilled in the art to which theinvention relates.

The present invention concerns a continuously variable transmission, ortransfer drive assembly, particularly suited for driving auxiliarydevices in an automotive vehicle. Of course, the principles of theinvention can be employed in a variety of applications wherecontinuously or infinitely variable speed ratios are desired.

In general terms, the invention provides a driving member assembly thatincorporates mechanical tensioning features to maintain proper tensionon a V-shaped belt driven by the rotating sheaves of the driving pulley.The driving member assembly also includes a disengagement mechanismoperable to isolate the belt from the rotation of the pulley sheaves. Inanother general aspect of the invention, the continuously variabletransfer drive assembly includes a driven member assembly that utilizesmechanical gearing to adjust the relative position between the rotatingsheaves of the driven pulley. In addition, the driven member assemblyincludes a fail-safe mechanism that automatically restores the drivenpulleys to a predetermined pitch or pulley ratio upon failure of powersupplied to the components of the driven member assembly.

With this general background, further details of the various embodimentsof the invention will be disclosed with specific reference to thefigures. Referring first to FIG. 2, the general components of thetransfer drive assembly 15 according to one embodiment is shown. Inparticular, the transfer drive assembly 15 includes a driving memberassembly 20 that is connected to a source of rotary power, such as aninternal combustion engine, and a driven member assembly 22, which isconnected to a driven device, such as an auxiliary device associatedwith a vehicle. In the illustrated embodiment, the driven memberassembly 22 can be connected to a cooling fan forming part of the enginecooling system. A continuous belt 24 is connected between the pulleys ofthe driving member assembly 20 and the driven member assembly 22. Thebelt 24 is preferably V-shaped and can be of a variety of knownconfigurations and materials. In the preferred embodiment, the belt 24is driven by frictional contact with the pulley of the driving memberassembly. Likewise, the driven member assembly 22 is propelled throughfrictional contact with the rotating belt.

In the present embodiment, the driving member assembly 20 includes adriving shaft 26 that can be configured to mount to the drive shaft ofthe engine or an auxiliary or PTO shaft driven by the automotive engine.The driven member assembly 22 can include a fan mounting cover 44 with apattern of screw bores 45 (FIG. 14) to which the engine cooling fan canbe engaged.

The present invention contemplates a conical pulley system engaged bythe continuous belt to transfer rotary power from the driving memberassembly 20 to the driven member assembly 22. Thus, the driving memberassembly 20 includes a rear sheave 28, having a conical engagementsurface 29, and a forward sheave 30, also having a conical engagementsurface 31. As is well known in the art, the two sheaves 28 and 30combine to form a pulley for driving the continuous belt 24. The V-shapeof the belt 24 conforms to the opposing conical surfaces 29 and 31 toprovide solid frictional contact during rotation of the driving memberassembly 20.

The driving member assembly 20 further includes a belt tensioningmechanism 32 that is preferably operably engaged to the rear sheave 28.The tensioning mechanism maintains tension in the rotating belt 24 byproviding pressure to the rear sheave 28. Pressure on the rear sheave 28pushes it toward the forward sheave 30 which consequently narrows thegap between the conical surfaces 29, 31. As this gap is narrowed, thecontinuous belt 24 is urged radially outward to thereby maintainappropriate tension on the belt.

For most pulley belt-driven automotive systems, the position of thedriving and driven pulleys is fixed to maintain appropriate tension inthe belt. However, with the use of a continuously variable system, thebelt 24 can be driven by or drive the appropriate pulleys at differingradii. Consequently, the belt tensioning mechanism 32 is important tomaintain proper belt tension, ensure efficient transfer of rotary motionbetween the two pulleys, and eliminate belt squeal associated with aloose or worn belt.

In a further feature of the driving member assembly 20, the pulleyformed by the rear sheave 28 and forward sheave 30 is permitted to slideaxially along the driving shaft 26. Changing the pulley ratio betweenthe driving member assembly 20 and driven member assembly 22 causes thecenterline of the belt 24 to shift axially relative to the driving shaft26. Thus, the pulley formed by the sheaves 28, 30 must be free to slideaxially to maintain proper alignment between the driving member pulleyand driven member pulley. Without this feature, the continuous belt 24will be skewed between the two pulleys, increasing belt wear and therisk of belt breakage. In the illustrated embodiment, the axial travelof the sheaves is limited at one end by the flange of the driving shaft26, and at an opposite end of the driving shaft 26 by a travel stop 34.

A second component of the continuously variable drive assembly 15 is thedriven member assembly 22. The assembly 22 can be fixed to the vehicle,preferably to the engine, by way of a mounting base plate 38. The drivenmember assembly 22 also defines a rotating pulley by the combination ofa rear sheave 40 and a forward sheave 42. As with the driving member,the two driven sheaves 40, 42 define conical engagement surfaces 41, 43,respectively. A fan mounting cover 44 is engaged to the forward sheave42 so that rotation of the pulley sheaves causes rotation of the cover44, and ultimately rotation of a fan attached to the cover.

In accordance with the preferred embodiment of the invention, thecontinuously variable ratio feature of the assembly 15 is accomplishedby a ratio adjustment mechanism 46 integrated into the driven memberassembly 22. In general terms, the adjustment mechanism 46 adjusts theposition of the rear sheave 40 relative to the forward sheave 42 toincrease or decrease the gap between the two sheaves. As explainedabove, moving the two sheaves together causes the belt 24 to be forcedradially outward to a larger driven radius. Similarly, moving the twosheaves apart allows the belt to drop deeper into the pulley groove, andtherefore run at a smaller driven radius. It is preferred that theadjustment mechanism 46 be associated with the driven pulley, ratherthan the drive pulley. However, a similar mechanism can be incorporatedinto the driving member assembly 20, or into both driving and drivenassemblies.

In a further feature of the preferred embodiment of the invention, thedriven member assembly 22 includes a fail-safe mechanism 48. In oneembodiment, the ratio adjustment mechanism 46 is powered by an electricmotor. When power is interrupted to the motor, the fail-safe mechanism48 forces the driven member assembly 22 to a predetermined pulley ratio.Details of the fail-safe mechanism 48 will be developed herein.

Referring now to FIGS. 3–12, specific features of the driving memberassembly 20 will be explained. The driving shaft 26 can include amounting flange 50 configured to engage a rotating shaft powered by thevehicle engine. The driving shaft 26 defines a splined shaft 51extending substantially along the length of the driving member assembly20. The travel stop 34 in the preferred embodiment can be a snap-ringfixed within a groove at the end of the splined shaft 51. At theopposite end of the shaft, and adjacent the mounting flange 50, thedriving shaft 26 defines a rear stop surface 52 which further limits theaxial travel of the rear and forward sheaves 40, 42. More specifically,the rear stop surface 52 is contacted by a floating sleeve 55 thatsupports the entire driven member assembly, including the pulley sheaves40, 42, on the driving shaft 26.

It is understood that the driving shaft 26 and its integral splinedshaft 51 are driven by a source of rotary motion. The rotation of thesplined shaft 51 is transmitted to the two pulley sheaves through thefloating sleeve 55. The floating sleeve includes inner splines 56 thatmate with the splined shaft 51. This splined interface between thefloating sleeve 55 and shaft 51 allows rotary motion to be transmittedbetween the two components, while permitting the floating sleeve toslide axially along the length of the shaft between the snap-ring 34 andrear stop surface 52.

At an end of the floating sleeve 55 adjacent the travel stop 34, thesleeve defines outer threads 57. These threads mate with correspondinginner threads 60 defined in the forward sheave 30. The outer threads 57and inner threads 60 are preferably machined threads so that the forwardsheave 30 can be firmly engaged, or fixed, to the forward end of thefloating sleeve 55. From the perspective of the floating sleeve 55, theforward sheave 30 is stationary, meaning that the sheave 30 cannot moveaxially relative to the sleeve. In contrast, the rear sheave 28 isarranged to slide axially relative to the sleeve 55.

The floating sleeve 55 also defines outer splines 58 situated beneaththe rear sheave 40. The rear sheave 28 then, also defines mating innersplines 62. Again, the splined interface between the floating sleeve 55and rear sheave 28 allows the sheave to translate axially along thesleeve, while rotary power is transmitted between the two components. Inthe preferred embodiment, a collar 63 is disposed around the outside ofthe rear sheave 28 adjacent the inner spline 62. In the illustratedembodiment, the rear sheave 28 is movable while the forward sheave 30 isrelatively stationary. It is understood, of course, that the roles ofthe two sheaves of the driving pulley can be reversed, with appropriatemodification to the other components of the driving member assembly 20.

In one feature of the invention, the driving member assembly 20 includesa disengagement mechanism 65 at the innermost radius of the pulleyformed by the rear sheave 28 and forward sheave 30. More specifically,the forward sheave 30 defines a bearing recess 61 (see FIG. 4), and therear sheave 28 defines a similar bearing recess 64 (see FIG. 5).Disposed within the forward bearing recess 61 is a front idler 66 andbearing 68. The front idler defines a conical surface 67. Likewise, therear bearing recess 64 receives a rear idler 69 supported by a rearbearing 71. The rear idler also defines a conical surface 70 so that thefront and rear idlers together define, in essence, a separate conicalpulley section.

Since the two idlers 66, 69 are supported relative to the correspondingsheaves 28, 30 by bearings, the pulley formed by the idlers isrotationally isolated from the pulley formed by the sheaves 28, 30. Inthe operation of the driving member assembly 20, as the drive assembly15 moves to a lower ratio, the belt 24 moves lower between the drivingmember sheaves. When the belt moves far enough, it contacts the conicalsurfaces 67, 70 of the idlers 66, 69, respectively, rather than thesurfaces of the primary sheaves 28, 30. When the belt is at thislocation, the rotation of the belt ceases since the idlers 66, 69 do notrotate with the rotating pulley sheaves. In this configuration, themechanism 65 completely disengages the driven member assembly 22, andconsequently the driven auxiliary device, from the rotary power source.In the case of a cooling fan, when the belt 24 reaches the disengagementmechanism 65, the rotation of the fan stops.

The driving member assembly 20 further includes a belt tensioningmechanism 32. Since the amount of belt tension required to prevent slipdepends on rotational speed, the mechanism 32 applies increasing axialforce to the belt as the speed increases. In accordance with a preferredembodiment of the invention, the belt tension is variable instead ofconstant, to increase the belt life and reduce component fatigue fromhigh belt loads. In other words, at lower rotational speeds, lower belttension is acceptable. Conversely, at higher speeds, higher belt tensionis necessary. Thus, the belt tensioning mechanism 32 is configured toprovide greater axial force at higher rotational speeds.

The inventive belt tensioning mechanism 32 contemplates two tensioningelements. The first element provides tensioning force as a function ofthe rotational speed of the driving member assembly 20. Specifically,this first element is a weight arm assembly 100. The weight arm assembly100 includes a number of weight arms 101 that are pivotally mounted tothe floating sleeve 55 at a pivot 102. As shown in more detail in FIG.8, the sleeve 55 defines a weight arm slot 103, with the pivot 102 atone end of the slot. The weight arm slot 103 provides clearance forpivoting of the weight arm 101.

The weight arm 101 carries a centrifugal weight 104 that is specificallysized to provide a predetermined axial force as a function of rotationalspeed. In one specific embodiment, the centrifugal weights 104 areformed of depleted uranium due to the high density of the material. In aspecific embodiment, the weight arm assembly 100 includes three weightarms 101 symmetrically disposed at 120° intervals around the floatingsleeve 55. At least three weights are preferred to avoid torsionalvibration problems. More weight arms and weights can be utilizedprovided they are symmetrically arranged around the floating sleeve 55.The magnitude of the centrifugal weights are calibrated based on themaximum required axial force and the centrifugal force generated byrotation of the weights. In the illustrated embodiment where theassembly drives an automotive cooling fan, the weights 104 can be about1–2 pounds.

It is understood that as the floating sleeve 55 rotates with drivingshaft 26, the weight arms 101 gradually pivot outward about pivot point102 due to centrifugal effects. As the weight arms 101 swing outward,they transmit an axial force to rear sheave 28 to push it closer to therelatively stationary forward sheave 30. This force transmission occursthrough a roller 107. More particularly, the roller 107 is affixed tothe rear sheave 28 through a roller bracket 106. The bracket is mountedto the rear-most surface of the rear sheave by a mounting screw 108engaged within screw bore 113 (see FIG. 6). The bracket 106 supports theroller 107 so that as the weight arm 101 presses against the roller,force is transmitted to push the rear sheave 28 axially.

The tension in the belt 24 tends to urge the belt deeper into the pulleygroove between the sheaves 28, 30. Thus, as the rotational speed of theshaft 26 decreases and the weight arms 101 decline, the belt will act topush the rear sheave 28 rearwardly to maintain constant pressure betweenweight arm 101 and the roller 107. In order to further help maintain theweight arm 101 in contact with the roller 107, a tether in the form ofan extension spring 110 is connected between the arm and a springbracket 109. The spring bracket is fixed to the rear sheave 28 beneaththe roller bracket 106 using the same mounting screw 108. In thespecific embodiment, the spring bracket 109 is partially disposed withina bracket recess 111 (see FIG. 6) to accommodate a reasonable length forthe extension spring 110. The tether or extension spring 110 constantlypulls the weight arm 101 back toward the roller 107. This preventsproblems with the driving member assembly 20 as it initially beginsrotating, when the weight arm would ordinarily be fully declined in theabsence of any centrifugal effects. Once the shaft 26 starts to rotate,however, the weight arms 101 would be flung outward, which can causedamage to the arms and rollers 107. The extension spring 110 eliminatesthis difficulty by keeping the idle position of the arms constrained.

Belt tension is not only a function of rotational speed, it is alsoaffected by the drive or pulley ratio—i.e., the ratio between thediameters of the driving and driven pulleys. In order to account forthis tensioning relationship, the belt tensioning mechanism 32 includesa second component in the form of a spring pack and lever system. Inaccordance with one embodiment of the invention, the floating sleeve 55is configured at its rear end into a number of spring guide blades 75,shown best in FIG. 7. In the illustrated embodiment, three such bladesare utilized. Each blade includes two bores through which a spring guide76 (FIG. 3) extends. An enlarged head 77 of the spring guides 76 preventtheir full passage through the blades 75. A compression spring 80 ismounted over each of the spring guides 76. In the illustratedembodiment, six such springs are utilized, two each for each guide blade75. The compression springs 80 are disposed between the floating sleeve55 and the rear sheave 28. Thus, the springs 80 maintain a continuouspressure against the rear sheave 28, regardless of the position of thebelt relative to the pulley sheaves.

However, it is well-known that the force supplied by a compressionspring is directly related to its displacement. Thus, when the rearsheave 28 is moved to its fullest rearward extent (to the left in FIG.3), the springs 80 generate their maximum restorative force. By the sametoken, when the rear sheave 28 is moved to its forward limit of travel,the springs 80 are only minimally depressed, so the force that theyapply is considerably weaker. When the belt is at its maximum radiallyoutward position, which can typically correspond with its highestrotational speed, the force being applied by the compression springs 80is at its lowest, which means that the spring pack is only minimallyeffective in maintaining tension in the belt 24.

In order to address this problem, a special lever system is incorporatedin one feature of the invention. With this feature, a spring plate 82 isslidably disposed over the rear sheave collar 63. The spring platedefines a spring bore 83, as depicted best in FIG. 9. A spring cup 84extends though each spring bore 83 and is held in position against therear surface of the spring plate 82. The compression spring 80 is thennested within each spring cup 84 so that the springs react against theguide blades 75 of the floating sleeve 55 to push forward against thespring plate 82.

Between each of the spring bores 83 is defined a roller support flange86. Each flange 86 supports a spring plate roller 87 engaged at pinbores 87 a. The spring plate 82 further defines a lever slot 88immediately adjacent or beneath each spring plate roller 87. The slots88 are defined to receive a lever arm 90 extending therethrough (seeFIG. 3). Each lever arm 90 is pivotally mounted to the floating sleeve55 at a pivot point 91. The pivot point is disposed within a lever slot95 (see FIGS. 7 and 8) so that the lever arm 90 has clearance to pivotrelative to the guide blades 75. The lever arm 90 includes a cam-edge 92that bears directly against the spring plate roller 87. The arm furtherincludes a lever arm roller 93 rotatably mounted at the end of the armopposite the pivot 91, as best shown in FIG. 3.

The lever arm roller 93 rides on a force transmitting surface 94 (seeFIGS. 3, 5, and 6) defined in the rear surface of the rear sheave 28. Itcan thus be appreciated that the force generated by the compressionspring 80 and reacted against the guide blades 75, is applied to thespring plate 82 by way of the spring cups 84. The spring plate 82 isurged forward (to the right in FIG. 3) so the spring plate roller 87contacts and pushes the lever arm 90. As the lever arm 90 is pushed,force is transmitted directly to the rear pulley sheave 28 through thelever arm roller 93.

In the other direction, as the rear sheave 28 moves rearward, or awayfrom the forward sheave 30, the lever arm 90 rotates about the pivotpoint 91. At the same time, the lever arm roller 93 rides radiallyoutwardly along the force transmitting surface 94. The cam-edge 92 thenpushes against the spring plate roller 87 to thereby translate thespring plate actually rearwardly (to the right). As the spring plate istranslated, the springs 80 are compressed even further.

In a further feature of the driving member assembly, the rear sheave 28includes a support hub 72. This support hub underlays the forward sheave30. When the rear sheave 28 is at its rearmost position, the support hub72 is exposed in the gap between the two sheaves, as best seen in FIG.12.

This action of the driving member assembly 20 is illustrated in thediagrams of FIGS. 11 and 12. In the configuration shown in FIG. 11, thedriving member assembly 20 is operating substantially at its maximumspeed. At this speed, the forward and rear sheaves are united and thesupport hub 72 is disposed fully underneath the forward sheave 30. Theweight arm 101 is at its greatest radial orientation and the lever arm90 is at the innermost end of the force transmitting surface 94.

As the speed of the rotational input decreases, the weight arms 101gradually recline, allowing the rear sheave 28 to translate axiallyrearward. As the rear sheave moves in that direction, it bears againstthe lever arm 90 causing the arm to rotate about its pivot point 91. Atthe same time, the lever arm, in particular the cam-edge 92, pushesagainst the spring plate roller 87, causing the spring plate 82 totranslate axially rearward. This movement compresses the springs 80 (notshown in FIG. 12).

In order to maintain a uniform force applied by the compression springs80, the cam-edge 92 of the lever arm 90 adopts a predefined curvature.In the specific embodiment, the curvature is a flattened S-shape asshown in FIG. 3. This curvature of the cam-edge 92 allows the springs 80to be pre-compressed to an axial force against the rear sheave 28sufficient to maintain proper belt tension even at the highest pulleyratios. At the same time, the configuration of the cam-edge 92 regulatesthe axial force transmitted to the rear sheave 28 as the compressionsprings 80 are depressed when the driving member assembly 20 is in theconfiguration shown in FIG. 12.

In the illustrated embodiment, the spring plate 82 provides a number ofspaced openings 89 between each of the roller support flanges 86. Theseopenings 89 are oriented for passage of each weight arm 101. As theconfiguration of the spring plate 82 illustrates, the weight arms areangularly offset from the spring pack portions of the assembly. In theillustrated embodiment, three weight arms are provided, requiring threeopenings 89 in the spring plate. Of course, additional weight arms canbe utilized. It is important, however, to have the arms orientedsymmetrically around the driving member assembly to avoid vibrationproblems associated with an eccentric weight.

An alternative embodiment of the driving member assembly is depicted inFIG. 13. In particular, the assembly 120 includes a driving shaft 121having a different configuration for mating with an output shaft of theengine. The assembly 120 includes a rear sheave 123 and a forward sheave124 that operates similar to the sheaves for the driving member assembly20. Both sheaves are supported on a floating sleeve 125 that is actuallymovable along the length of the shaft 121. The driving member assemblycan also include a disengagement mechanism 126 similar to the mechanism65 described above. Likewise, the assembly 120 can include a weight armassembly 127 that centrifugally tightens the belt riding between thesheaves 123, 124.

In one modification from the prior embodiment, the floating sleeve 125supports a spring guide 132 onto which a compression spring 131 ismounted. The rear sheave 123 defines a spring recess 130 in line withthe spring guide 132. The compression spring is then engaged within therecess so that it provides outward forces against the floating sleeve125 and directly against the rear sheave 123. In this configuration, thelever arm 90 of the prior embodiment is eliminated.

In place of the lever arm, the weight arm assembly 127 includes aspecially configured weight arm 133. Specifically, the weight armdefines a cam-edge 134 that bears against a roller 135 supported on therear sheave 123. The cam-edge 134 follows a specific configuration tooptimize the axial force applied to the rear sheave 123 at the higherrotational speeds. The cam-edge 134 of the weight arm 133 follows ageometry similar to the cam-edge 92 of the lever arm 90 in the previousembodiment. In both cases, appropriate tensioning force is maintainedthroughout the range of rotational speeds.

Details of the driven member assembly 22 are depicted in FIGS. 14–20. Asexpressed above, the driven member assembly includes a ratio adjustmentmechanism 46 that operates on a movable rear sheave 40. In addition, thedriven member assembly includes a fail-safe mechanism 48 that isintegrated with the ratio adjustment mechanism 46 to account for a lossof power to the ratio adjustment mechanism. In accordance with apreferred embodiment of the invention, the adjustment mechanism is motordriven. Thus a loss of electrical power to the motor can causedifficulties with respect to the pulley ratio in the absence of afail-safe mechanism.

Turning to FIG. 14, it can be seen that the forward sheave 42 isrotatably supported on a needle/thrust bearing 140. An oil seal 141 isalso provided between the rotating sheave and non-rotating components ofthe driven member assembly 22. Likewise, the rear sheave 40 is supportedon a combination needle/thrust bearing 142. A rotating seal 143 is alsoprovided between rotating rear sheave 40 and the stationary elements ofthe driven member assembly.

In one feature of the driven member assembly, the rear sheave 40 isinterlocked with the forward sheave 42 so that both components rotatetogether. In order to accomplish the ratio adjustment feature, however,the rear sheave 40 must be permitted to move axially with respect to therelatively stationary forward sheave 42. Thus, in the illustratedembodiment the forward sheave 42 is provided with a number of slots 144.The rear sheave 40 includes a like number of interlocking prongs 145. Apreferred arrangement of the slots and prongs is depicted in the endview of the rear sheave 40 shown in FIG. 16. It can be seen that theinterlocking slots and prongs 144, 145 are arc segments. In the specificembodiment, six such interlocking components are provided to adequatelytransfer torque between the two components and maintain their unisonrotational operation. The prongs 145 are configured to readily slideaxially along the length of a corresponding slot 144.

The ratio adjustment mechanism 46 relies upon the application of amechanical force against the rear sheave 40 to move it closer to orfurther away from the forward sheave 42. In the preferred embodiment,the adjustment mechanism 46 includes a bearing pressure plate 148 thatis at least partially disposed within the rear sheave 40. The bearingpressure plate 148 directly contacts and presses against the bearing 142that rotationally supports the rear sheave 40. The adjustment mechanism46 further includes a number of force pins 149 that press against thebearing pressure plate 148. The force pins 149 are supported by apressure plate 152.

In the preferred embodiment, as shown in FIG. 17, the pressure plate 152includes a plurality of radially extending spokes 153. A force pin 149is connected at the end of each of the spokes 153. Preferably, six suchspokes are provided, along with corresponding force pins, uniformlydispersed around the circumference of the pressure plate 152. In thisway, pressure applied by the force pins 149 is evenly distributedagainst the bearing pressure plate 148.

Movement of the pressure plate 152 is accomplished by operation of anactuation screw 154. Specifically, the actuation screw 154 includes anenlarged head 155 that bears against the pressure plate 152 through athrust bearing. The opposite end of the screw 154 defines a screwthreaded portion 156. The threaded portion 156 is configured tothreadedly engage internal screw threads 162 of a split nut 158. In theillustrated embodiment, the split nut is disposed beneath the forwardsheave 42.

In operation, the actuation screw 154 is rotated so that the threadedportion 156 is threaded into the split nut 158. As the actuation screw154 is continuously threaded, the head 155 bears against the pressureplate 152, which causes the force pins 149 to push against the bearingpressure plate 148. Continued rotation of actuation screw 154 ultimatelycauses the rear sheave 40 to be pushed closer to the forward sheave 42.As indicated above, moving the two sheaves together pushes their conicalsurfaces 41 and 43 against the V-shaped belt 24 pushing it radiallyoutward to thereby change the pulley ratio.

In order for the actuation screw 154 to accomplish its appointedfunction, the split nut 158 must be held axially stationary relative tothe rear sheave 40. Thus, the split nut 158 is mounted within a splitnut holder 159. A retainer 160 is internally threaded into the split nutholder 159 to trap the split nut 158 between the holder and theretainer. The split nut holder 159 is itself threaded into a supportshaft 164 at a threaded engagement 165. The support shaft 164 is mountedto the base plate 38, and is therefore stationary with respect to theratio adjustment mechanism 46.

Referring to FIG. 18, it can be seen that the interior of the supportshaft 164 is configured into an array of pin channels 166. These pinchannels are aligned with each of the force pins 149 and with the spokes153 of the pressure plate 152. In this way, the pressure plate 152 isprevented from rotating, its movement being limited to axialdisplacement along the pin channels 166 of the support shaft 164.

As expressed above, the ratio adjustment mechanism 46 is driven by amotor. In the illustrated embodiment, a motor 170 is mounted on themounting plate 38 by a mounting bracket 169 (FIG. 15). The motor ispreferably an electric motor driven by the vehicle electrical system. Ina most preferred embodiment, the motor 170 is driven by signals from anengine control module that monitors the engine operation andperformance. Specifically, the engine control module can makedeterminations as to when the transfer drive assembly ratio must bechanged and to what extent. Consequently, the motor 170 must be capableof intermittent action and incremental motion. Preferably, the motor 170is a gear motor driven by a PWM controller, although other motors, suchas a stepping motor, can be used. In one specific embodiment, the motoris a model IM-15 motor provided by Globe Motors Co.

The motor 170 drives a worm 171 which mates with a worm gear 172. In theillustrated embodiment, the motor is oriented transverse orperpendicular to the axis B of the driven member assembly 22. Thus, theworm and worm gear combination transmits the rotary power of the motorto rotational movement of the worm gear 172. It is understood, however,that other motor and gearing combinations are contemplated by thepresent invention. For instance, a rack and pinion arrangement can beutilized to translate power from a linear motor to rotational movement.

The worm gear 172 is mounted to a worm gear shaft 173. The worm gearshaft 173 passes through a hollow end of the actuation screw 154. Theworm gear shaft 173 is supported at an opposite end by a thrust bushing174 mounted within the mounting base plate 38.

The actuation screw 154 defines a pair of opposite engagement slots 175.A dowel pin 176 passes through the worm gear shaft 173 and is orientedwithin the engagement slots 175. In this manner, the worm gear shaft 173can transmit rotational movement to the actuation screw 154 by way ofthe dowel pin 176. At the same time, the actuation screw 154 is free toslide axially along the axis B with the dowel pin 176 sliding along theengagement slots 175. It can therefore be appreciated that rotation ofthe worm gear shaft 173 under power from the motor 170 causes directrotation of the actuation screw 154.

When the motor 170 directs rotation of the worm gear shaft 173 in onedirection, the actuation screw 154 is threaded deeper into split nut158. As the actuation screw 154 is threaded into the nut it advancestoward the rear sheave 40, pushing the rear sheave as described above.In the alternative, rotation of the motor 170 in the opposite directioncauses the actuation screw 154 to unthread from the split nut 158. Asthe actuation screw 154 is retracted, the bearing pressure plate 148moves away from the bearing 142 supporting the rear sheave 40. Thetension within the rotating drive belt 142 causes the belt to projectdeeper into the gap between the rear and forward sheaves, therebypushing the rear sheave 40 back toward the pressure plate 152. Thus, thebearing pressure plate 148 is always substantially in contact with theneedle/thrust bearing 142 of the rear sheave 40.

The driven member assembly 22 further includes a fail-safe feature thataccounts for a loss of electrical power to the ratio adjustmentmechanism 46. In the preferred embodiment, this mechanism 48 includes asolenoid 180 mounted to the free end of the support shaft 164. Morespecifically, the solenoid 180 is supported by a mounting bracket 182 onthe split nut holder 159. A number of control wires 181 electricallyconnect the solenoid 180 to an external electrical source. Since thesupport shaft 164 is stationary, the control wires can pass along achannel defined in the shaft, exiting adjacent the mounting base plate38. The solenoid 180 is preferably electrically connected to the vehicleelectrical system, and most preferably to the engine control module.Thus, when power is interrupted to the adjustment mechanism motor 170,power is also interrupted to the solenoid 180. In one specificembodiment, the solenoid 180 can be a low profile push-pull solenoid,such as a model 129415-023 solenoid provided by Lucas Ledex Co.

The solenoid 180 includes a solenoid shaft 183 that is held in itsactuated position as long as power is provided to the solenoid 180. Thesolenoid shaft 183 is threadedly engaged to an engagement finger holder185. This finger holder supports a number of engagement fingers 186 thatproject toward the split nut 158. More particularly, the engage fingers186 contact a control ramp surface 161 of the split nut 158.

Operation of the engagement fingers is best understood following anexplanation of the structure of the split nut 158, with specificreference to FIGS. 19 and 20. The split nut 158 includes a number ofseparable components 158 a–158 c. When the components are combined, theydefine the internal screw threads 162 that are engaged by the threadedportion 156 of the actuation screw 154. However, when the components ofthe split nut are separated, the internal screw threads 162 areinterrupted and the threaded portion 156 of the actuation screw 154 hasno screw threads to engage. The component 158 a–158 c are separated by asplit gap 195. Preferably, this gap is zero when the components of thesplit nut are combined. On the other hand, when the split nut isseparated, this gap 159 is large enough so that the internal threads ofthe split nut cannot contact the threaded portion 156 of the actuationscrew 154.

In order to maintain the integrity of the split nut 158 and insurerepeatable separation and combination of its components 158 a–158 c, thesplit nut includes a number of guide tabs 196 projecting therefrom.These guide tabs are aligned to slide within corresponding guide slots197 defined in the retainer 160 (see FIG. 21). The retainer 160 alsoincludes a number of finger bores 198 aligned with the engagement fingerholder 185 to receive the engagement fingers 186 therethrough.

With this background on the split nut 157, the operation of theengagement fingers 186 can be more readily understood. As the engagementfingers 186 are pushed rearward, i.e. toward the split nut 158, thefingers contact the control ramps 161 of each of the split nutcomponents 158 a–158 c. As the fingers 186 move along the ramp, theycontinue until they reach the outer diameter of the split nut 158. Atthis point, the split gaps 195 are essentially closed and the internalscrew threads 162 of the split nut are defined.

On the other hand, with the engagement fingers 186 are retracted, theymove away from the control ramps 161. Once the fingers have cleared theramps and are no longer in contact with the split nut, the components158 a–158 c are free to separate. The overall integrity of the split nut158 is maintained by the tabs 196 sliding along the slots 197. Theseparation of the split nut components 158 a–158 c can be accomplishedby separation springs 199 mounted within the split nut. The separationsprings can be compression springs or leaf springs supported within eachcomponent to span the split gaps 195.

During normal operation, the solenoid 180 is powered and the solenoidshaft 183 is maintained in its actuated position. However, when power isremoved from the solenoid, the shaft 183 is pushed away from theretainer 160 by operation of a return spring 187. As shown in FIG. 14,the return spring is contained within the engagement finger holder 185and the retainer 160. Thus, the return spring 187 in essence pushes theengagement fingers 186 away from the split nut 158, allowing itscomponents to separate.

When the split nut 158 is separated, the threaded portion 156 of theactuation screw 154 no longer has a threaded reaction surface to operateagainst. In this event, the fail-safe mechanism 48 provides means forpushing the rear sheave 40 forward to the forward sheave 42, therebyincreasing the pulley ratio. This action is accomplished by a returnspring 190 disposed within the support shaft 164. The return spring 190is situated between a spring carrier 191 at one end and a reactionflange 192 internally formed within the support shaft 164. The springcarrier 191 is retained relative to the actuation screw 154 by way of acarrier nut 193. The large return spring 190 can exert force on thespring carrier 191 through a thrust bearing 194 that can be provided toreduce rotational drag on the actuation screw.

The fail-safe mechanism 48 of the present invention is operable toreturn the driven member assembly to a predetermined pulley ratio. Forthe purposes of explanation, the illustrated embodiment provides afail-safe ratio of 1:1. When the split nut components 158 a–158 c areseparated, the response of the fail-safe components depends upon thecurrent pulley ratio. For a ratio greater than the predetermined value(1:1 in the present example), the mechanism 48 drives the rear sheave 40forward. For ratios less than the predetermined value, the mechanismallows the belt tension to separate the two sheaves.

Looking first at a pulley ratio greater than the specific 1:1 value, therear sheave 40 is separated from the forward sheave 42. When the splitnut components 158 a–158 c separate, the threaded portion 156 of theactuation screw is free to slide axially forward along the axis B. Thelarge return spring 190 pushes the spring carrier 191 forward, whichcontacts the carrier nut 193 to further push the actuation screw 154forward. As the actuation screw 154 is pushed forward, the enlarged head155 contacts the pressure plate 152, causing the force pins 149 to bearagainst the bearing pressure plate 148. The bearing pressure plate 148pushes against the rear sheave 40 until the spring carrier 191 reachesits limit of movement, at which point the rear sheave is immediatelyadjacent the forward sheave 42. In a specific embodiment, the twosheaves are separated by a gap of about 0.5 inches at their closestpoint.

The large return spring 190 is calibrated to provide sufficient force toact against the operating tension in the belt 24. Moreover, the forwardmovement of the rear sheave is limited by the movement of the springcarrier 191 as the large spring 190 extends. Specifically, in thepreferred embodiment, the spring carrier butts against the split nutholder 159 to limit its axial movement. The position of the rear sheave40 is thus fixed once the carrier contacts the nut holder, which therebyestablishes the predetermined pulley ratio.

When the pulley ratio is less than the predetermined value (1:1), thespring carrier 191 is already in contact with the nut holder 159, thethreaded portion 156 of the actuation screw 154 extends deeply into thenut holder, and the carrier nut 193 is disposed apart from the springcarrier. When the split nut components 158 a–158 c separate, thethreaded portion 156 is released and the actuation screw 154 is freelyto move axially rearward. The belt tension is then free to push the rearsheave 40 away from the forward sheave 42. As the rear sheave movesback, the bearing pressure plate 148 pushes against the force pins 149,which push against the pressure plate 152, and ultimately against theenlarged head 155 of the actuation screw. As the actuation screw 154 ispushed rearward, the carrier nut 193 moves into contact with the springcarrier 191 which further compresses the large spring 190. Thisrestorative movement continues until the force generated by the largespring 190 matches the force created by the belt tension. At this point,the driven pulley is at the predetermined ratio.

The driven member assembly 22 is indicative of one embodiment of thetransfer drive assembly according to the present invention. Anadditional embodiment is illustrated in FIG. 22. Specifically, a drivenmember assembly 200 includes a rear sheave 201 and forward sheave 202.In this instance, the fan mounting flange 204 is engaged at one end of adriven shaft 205. The froward sheave 202 is mounted at the opposite endof the driven shaft 205. The driven shaft 205 is rotatably supported bya bearing housing 208 by way of a pair of tapered roller bearings 209.This bearing housing 208 can be mounted to the vehicle or engine.

A screw flange 212 is mounted to the bearing housing 208. The flange 212defines external screw threads that mate with corresponding threads 215on a thrust collar 214. The thrust collar applies force against theforward sheave 201 through a needle bearing 216.

The ratio adjustment mechanism includes a motor 220 that is arrangedparallel to the axis of the driven shaft 205. This configuration for themotor allows the driven member assembly 200 to be mounted within avehicle having particular space requirements. The motor 220 drives apinion gear 219 which engages a spur gear 218. A spur gear 218 isattached to the thrust collar 214. Thus, rotation of the pinion gear 219by the motor 220 is translated to rotation of the spur gear 218. As thespur gear rotates, so does the thrust collar 214. Rotation of the thrustcollar 214 causes its internal threads 215 to advance or retract alongthe external threads 213 of the screw flange 212. In this way, theposition of the rear sheave 201 relative to the forward sheave 202 canbe modified to adjust the pulley drive ratio.

In an alternative embodiment of the driving member assembly, an assembly230 shown in FIG. 23 includes a driving shaft 232. The assembly includesa rear sheave 234 and a forward sheave 235. A disengagement mechanism236 can be disposed between the two sheaves, as with prior embodiments.

The driving member assembly 230 provides a different tensioningmechanism 238 than with the prior embodiments. In particular, themechanism 238 includes a compression spring 240 that reacts between thedriving shaft 232 and a spring cup 241. A force transfer lever 243 ispivotally mounted at one end to the driving shaft 232. A transfer roller244 is provided at the opposite end of the transfer lever 243. Thespring cup 241 includes opposite rollers 246 that rotate along thetransfer lever 243.

In operation of this embodiment of the driving member assembly 230, asthe rear sheave 234 moves rearward, it exerts pressure against thetransfer roller 244. This pressure cause the transfer lever 243 to pivotradially outward relative to the driving shaft 232. As the transferlever pivots outward, the rollers 246 of the spring cup roll along thelever, causing the spring cup 241 to be displaced axially andrearwardly. As the spring cup moves rearwardly, the compression spring240 increases its resistant force until equilibrium is established. Whenviewed in a different sense, the compression spring 240 transfers atensioning force through the spring cup 241 to the transfer levers 243,through the rollers 244 and against the rear sheave 234 to push ittoward the forward sheave 235.

While the invention has been illustrated and described in detail in thedrawings and foregoing description, the same is to be considered asillustrative and not restrictive in character. It should be understoodthat only the preferred embodiments have been shown and described andthat all changes and modifications that come within the spirit of theinvention are desired to be protected. For example, in the depictedembodiments, the rear sheave of the drive assembly is movable withrespect to the relatively stationary forward sheave. This arrangementcan be reversed with appropriate modification to the inventive elementsof the system.

For instance, in some embodiments, the weight arm assembly, such asassembly 100, can be mounted differently. In one modification, theweight arms 101 can be pivotably mounted to the rear sheave 40 itself,rather than to the floating sleeve.

In alternative embodiments, certain of the features described above canbe eliminated. For instance, the disengagement mechanism, such asmechanism 65, need not be incorporated into all variable ratio transferdrive assembly designs. Likewise, a transfer drive assembly canincorporate several of the aforementioned inventive features, whileeliminating the weight arm assembly and/or other components of thetensioning mechanism, such as mechanism 32. Moreover, other tensioningsystems can be substituted for certain specific embodiments.

1. A variable ratio drive system (15) connectable between a source of rotary motion and a driven device, said system comprising: a driving member (20) having a rotating shaft (55) connectable to said source of rotary motion for rotation about a drive axis (A); a driven member (22) connectable to said driven device; and a belt (24) connected between said driving member (20) and said driven member (22) and operable to transmit rotary motion therebetween; wherein said driving member (20) includes; a first sheave (30) and a second sheave (28), each having a conical surface (31, 29) configured for frictional engagement with the belt (24); means (57, 60) for connecting said first sheave (30) to said rotating shaft (55) for rotation therewith; means for connecting said second sheave (28) to said rotating shaft (55) for rotation therewith while permitting translation relative thereto along said drive axis (A); a weight arm assembly (100) pivotally mounted to said rotating shaft (55) to pivot centrifugally outward relative to said drive axis (A); at least one force transmitting member disposed between said weight arm assembly (100) and said second sheave (28) to apply force to said second sheave (28) as a function of the centrifugally outward pivoting of said weight arm assembly (100); at least one spring (80) disposed between said rotating shaft (55) and said second sheave (28) to apply force to said second sheave (28) so as to help maintain tension in said belt (24); and a lever system pivotally mounted between said rotating shaft (55) and said second sheave (28) to apply force to said second sheave (28) so as to also help maintain tension in said belt (24).
 2. The variable ratio drive system according to claim 1, wherein at least one said force transmitting member includes a roller (107) mounted on said second sheave (28) and arranged to contact said weight arm assembly (100) as said assembly (100) pivots centrifugally outward.
 3. The variable ratio drive system according to claim 2, wherein said weight arm assembly (100) includes: at least one arm (101) pivotably mounted at a first end (102); and a weight (104) attached to each said arm (101) at an opposite second end.
 4. The variable ratio drive belt system according to claim 3, wherein at least one said arm (133) includes a substantially S-shaped portion (134) between said first and second ends that contacts at least one said roller (135), and said S-shaped portion includes both a concave cam-edge and a convex cam-edge.
 5. A pulley apparatus for a belt drive system (15), said pulley apparatus comprising: a first sheave (30) defining a first conical surface (31); a second sheave (28) defining a second conical surface (29) opposite said first conical surface (31) for engaging a belt (24) therebetween; a shaft (55) engaged to at least one of said first sheave (30) and said second sheave (28) for rotation therewith; an idler element (65) disposed between said first sheave (30) and said second sheave (28) and defining two conical surfaces (67, 70) situated opposite each other, said two conical surfaces (67, 70) having respective surface areas that are substantial commensurate with each other, and said two conical surfaces (67, 70) being substantially aligned with said first and second conical surfaces (31, 29) and configured for engaging a belt (24) therebetween; and means for supporting (68, 71) said idler element (65) relative to said first and second sheaves (30, 28) to isolate said idler element (65) from rotation with said first and second sheaves (30, 28).
 6. The pulley apparatus according to claim 5, wherein said means for supporting (68, 71) said idler element (65) includes a thrust roller bearing.
 7. The pulley apparatus according to claim 5, wherein: said second sheave (28) is movable along an axis (A) of said shaft (55) relative to said first sheave (30); and said idler element (65) includes a first idler component (66) mounted within said first sheave (30) and a second idler component (69) mounted within said second sheave (28), one of said first and second idler components (66, 69) defining a hub (72) arranged to span the distance between said first sheave (30) and said second sheave (28) when said sheaves (30, 28) move apart.
 8. The pulley apparatus according to claim 5, wherein the sidewall heights of said two conical surfaces (67, 70) defined by said idler element (65) are respectively greater than the sidewall thickness of said belt (24). 